1. Field of the Invention
The present invention relates to a hermetic rotary compressor and, in particular, to a hermetic rotary compressor which is capable of improving the effect of reducing noise due to pressure pulsation generated during a gas suction and discharge process, and at the same time improving the compressing efficiency of the compressor by reducing compressive driving force.
2. Description of the Prior Art
Generally, a rotary compressor is an apparatus for compressing gas, and there are many kinds of compressors depending on its method of compressing the gas including a rotary compressor, a reciprocating compressor, a scroll compressor, etc.
Each of these compressor includes a hermetic vessel having a certain space portion, an motor unit mounted on the hermetic vessel for thereby generating driving force, and a compression unit which receives the driving force from the motor unit for thereby compressing gas.
As an example of the above-mentioned compressors, a hermetic rotary compressor will be described as follows with reference to FIGS. 1 and 2.
FIG. 1 is a front cross-sectional view illustrating a general rotary compressor, and FIG. 2 is a horizontal cross-sectional view illustrating a general rotary compressor.
As illustrated therein, the motor unit is mounted on one side portion of the hermetic vessel 1, and the compression unit is mounted on the other side portion of the hermetic vessel at a certain distance from the motor unit.
The motor unit includes a stator 2 fixedly connected to the inner surface of the hermetic vessel 1, and a rotator 3 connected to be rotatable in the stator 2.
And, the compression unit includes a crankshaft 4 which is press-fitted to the inner diameter of the rotator 3 and has an eccentric portion 4a formed at one end of the crankshaft 4, and a cylinder 5 in which the eccentric portion 4a of the shaft 4 is inserted into a space portion 11 at which gas is sucked and compressed are mounted on the hermetic vessel.
In addition, the compression unit includes upper and lower bearings 7 and 8 which is bolted to the upper and lower surfaces of the cylinder 5 for thereby supporting the crankshaft 4 and enclosing the space portion 11 of the cylinder 5, a rolling piston positioned in the space portion 11 of the cylinder 5, revolving according to the rotation of the crankshaft 4, an eccentric portion 4a of the crankshaft 4 being inserted into the rolling piston 9, a vane 10 which is inserted into one side of the cylinder 5 in order to linearly reciprocate in a radius direction of the cylinder 5 as one end of the vane 10 contacts the outer surface of the rolling piston 9 during the rotation of the rolling piston 9, whereby the space portion formed by the inner surface of the cylinder 5 and the outer surface of the rolling piston 9 is partitioned into a suction area 11a and a compression area 11b.
And, a suction hole 5a through which gas is sucked into the cylinder 5 is formed in the suction area 11a of the cylinder 5, more specifically, at one side of the cylinder 5 neighboring the vane 10. A discharge port 5b through which compressed gas is discharged is formed in the compression area 11b of the cylinder 5, that is, at the other side of the cylinder 5 neighboring the vane 10. The above discharge port 5b is communicated with a discharge hole 7a formed at the upper bearing 7, and the discharge hole 7a can be formed at the lower bearing 8 connected to the lower surface of the cylinder 5.
A inlet pipe 12 through which gas is sucked is connected to a side wall of the hermetic vessel 1, a outlet pipe 13 through which gas is discharged is connected to the upper side of the hermetic vessel 1, and oil(not shown) is filled in the bottom of the hermetic vessel 1.
In the drawings, reference numeral 14 denotes a discharge valve, 15 denotes a retainer, 16 denotes a muffler, and 17 denotes an accumulator.
The operation of the above general hermetic rotary compressor will be described as follows.
When the crankshaft 4 is rotated by an applied current, along with the rotator 3, the rolling piston 9 connected to the eccentric portion 4a of the crankshaft 4 is revolved around the crankshaft 4 in the cylinder space portion 11 while being in contact with the vane 10.
Due to the volume change of the space portion 11 formed by the inner surface of the cylinder 5 and the outer surface of the rolling piston 9 by the revolution of the rolling piston 9, a gaseous refrigerant of low temperature and pressure is sucked into the space portion 11 of the cylinder 5 through the inlet pipe 12 and the suction hole 5a to thereafter be compressed into gas of high temperature and pressure, and the compressed gaseous refrigerant of high temperature and pressure is discharged through the discharge port 5b, the discharge hole 7a, and the discharge valve 14.
Herein, the process in which gaseous refrigerant is sucked, compressed, and then discharged according to the rotation of the crankshaft 4 will be described in more detail with reference to FIGS. 3, 4, and 5.
FIGS. 3, 4, and 5 are horizontal cross-sectional views illustrating the operational process of the rotary compressor.
First, as shown in FIG. 3, when the semimajor axial front end (A) of the eccentric portion 4a of the crankshaft 4 is in contact with the vane 10, the discharge stroke is terminated and at the same time the suction stroke is terminated.
And, as the crankshaft 4 is rotated, and thereby the space portion 11 is converted to the suction area 11a and the compression area 11b by the vane 10 at a position that the semimajor axial front end of the eccentric portion 4a is displaced from the vane by 180 degrees as illustrated in FIG. 4, gaseous refrigerant is sucked into the suction area 11a and at the same time the volume of the compression area 11a is reduced, whereby the gas is progressively compressed.
And, when the crankshaft 4 is rotated, and thereby the semimajor axial front end of the eccentric portion 4a passes an angle of 180 degrees and then moves to the discharge port 5b, the amount of gaseous refrigerant sucked into the suction area 11a and the pressure of the compression area 11b is increased at the same time, whereby the pressure of the compression area 11b becomes higher compared to discharged gas. In this case, the discharge valve 14 is opened, and compressed gas is discharged through the discharge port 5b and the discharge hole 7a.
Meantime, when the rolling piston 9 continues to repeat the process of sucking, compressing, and discharging gaseous refrigerant while revolving during the operation of the above compressor, noises due to pressure pulsation are generated. In this regard, many studies for reducing noises due to pressure pulsation is in progress in order to obtain an resonance effect at the space portion 11 of the cylinder 5.
With reference to FIGS. 6 and 7 illustrating an embodiment of a conventional noise reduction structure in order to reduce the above-mentioned pressure pulsation, a surge recess 18, an unpierced hole having a certain diameter and depth, is formed between 150 and 270 degrees from the vane 10 in a rotational direction of the crankshaft 4.
With respect to the position at which the above surge recess 18 is formed, there arises a malfunction that compressed gas flows back to the suction side at every angles at which the surge recess 18 is formed. When the angle is increased, the loss of re-expansion is increased as much, while the compression work(compressive driving force) of the compressor according to the surge recess 18 is decreased, thereby obtaining a gain of compressive driving force.
In regard to compression efficiency, when the performance of the compressor is analyzed based on a P-V diagram in FIG. 8, there arises a difference between a re-expansion loss and a compressive driving force gain within the space portion according to each position of the crankshaft in the process of the compression stroke during a single rotation of the crankshaft.
That is, it is shown that if the rolling piston 9 is positioned at 24 degrees from the vane 10, the re-expansion loss and the compressive volume gain or compressive driving force gain are small, if positioned at 90 degrees, the compressive volume gain of gas to be compressed becomes larger than the re-expansion loss thereof, and if positioned at 160 degrees, the compressive volume gain of gas to be compressed becomes smaller than the re-expansion loss thereof.
However, in the above-described conventional noise reduction structure, a simple tubular type unpierced hole is formed, so that noise reduction using resonance effect is not enough. Also, the unpierced hole is placed at a position of a high compressed state during the compression, thereby causing a re-expansion loss.
In addition, the conventional noise reduction structure is a certain set range considering only the discharge side with regard to pulsation noise reduction, rather than a proper range considering compressing efficiency as well.
Therefore, considering the above description, in the conventional rotary compressor, there is a problem that the surge recess for reducing noises due to pressure pulsation cannot maximize noise reduction, and the compressing efficiency is reduced.